Thermal energy storage devices (or thermal stores) are used to receive and then store heat, cold, or thermal energy for a period of time until it is needed for a useful purpose. Such thermal storage devices allow high or low (including cryogenic) temperature energy to be temporarily stored for later use, and offer the possibility of storing thermal energy, for example, for later conversion to electricity, for use in an air liquefaction process to reduce the energy consumed by the process, or to provide cooling for co-located processes. It is known to store thermal energy by increasing or decreasing the temperature of a substance, by changing the state (i.e. solid, liquid, or gas) of a substance, or by a combination of the two.
A thermal store typically operates on a three phase thermal storage process: charge, store and discharge. FIG. 1 shows a representation of a generic thermal store. The store includes a thermal mass 10, passages 20, surface features 30, an inlet 40 and an outlet 50. In a charge phase a heat transfer fluid (liquid or gas), hereinafter referred to as HTF, is passed through the inlet 40, into the passages 20, and out of the outlet 50, to either heat or cool the thermal mass 10. In a storage phase the thermal energy is then stored in the thermal mass 10 until required. In a discharge phase, the HTF is passed through the inlet 40, into the passages 20, and out of the outlet 50, over the thermal mass 10 to recover the thermal energy for transfer to another process. The thermal mass 10 includes surface features 30 to promote heat transfer. In known thermal stores the charge and discharge phases are symmetric, i.e. the HTF flow rates are the same during the charge and discharge phases.
One application in which such thermal stores are used is in the field of storing energy for generating electricity. An effective means of storing energy on a large scale is required to balance consumer demand for electricity with generating capacity, and to smooth out levels of intermittent supply from, for example, renewable energy sources. Energy demands vary on hourly, daily, weekly and seasonal bases. Alternative solutions for balancing supply and demand other than using traditional fossil fuel plants are now recognised as important to help control greenhouse gas emissions.
WO 2007/096656 discloses a cryogenic energy storage system which exploits the temperature and phase differential between low temperature liquid air and ambient air, or waste heat, to store energy at periods of low demand and/or excess production, allowing this stored energy to be released later to generate electricity during periods of high demand and/or constrained output. The system comprises a means for liquefying air during periods of low electricity demand, a means for storing the liquid air produced and an expansion turbine for expanding the liquid air. The expansion turbine is connected to a generator to generate electricity when required to meet shortfalls between supply and demand.
A major constraint to the efficiency of such cryogenic energy storage systems is the amount of cold energy remaining in the process air exhausted from the expansion turbine following expansion. The combination of cryogenic power storage and thermal energy storage provide a means of matching the electricity supply to meet variations in supply and demand. In particular, a thermal store can be used with a cryogenic energy storage system to recover and store the cold energy released when the cryogenic energy store is generating power and release the cold energy to reduce the energy cost of manufacturing cryogenic fluid when recharging the cryogenic energy storage system. FIG. 2 shows an example of a cryogenic energy storage system including an air liquefier module 60, a cryogenic liquid store 70, a cryogenic power recovery module 80, and a thermal store 90. FIG. 2 shows that there is a need for thermal storage because of the required time shifting between the generation of electricity and the need to generate further cryogen.
Therefore, there is a need for an efficient cold energy thermal store which facilitates the recovery of cold energy from exhaust gas, and which allows cold energy to be stored as high grade cold, to maximise the energy efficiency of the later recovery for use during the liquefaction phase, facilitating the production of more cryogen, and/or to provide cooling for co-located processes.
When a thermal store is used to store cold energy in a cryogenic energy storage system, the discharging phase and charging phase of the thermal store may be of different durations. In particular, the discharging phase is typically four or more times longer than the charging phase. Because of the mismatch between the periods of high and low demand and the different duration charging and discharging phases, there is a need for a flexible thermal storage system that can be charged and discharged at different rates. The need for such a thermal storage system presents a generic need for a thermal store in which the heat recovery, HTF pressure loss, and storage capacity can be optimised for an ‘asymmetric’ operating regime, i.e. a regime in which the charging and discharging of the thermal store are not carried out at the same HTF flow rates.
These needs also apply to systems in which thermal energy is stored as heat. Therefore, there is a need for an efficient thermal store which facilitates the recovery of heat, cold or thermal energy with high energy efficiency from exhaust gas.
Theory of Packed Bed Thermal Storage
The inventors have realised that it is important to optimise the design of the interface between the HTF and the thermal mass to ensure there is good transfer of heat from the HTF to the thermal store at a low pressure drop. In general it is desirable to provide a solid thermal mass having a large surface area, and features to break up the thermal boundary layer at the HTF-solid interface, in order to promote optimum heat transfer. However, such features increase the friction between the solid thermal mass and the HTF and hence increase pressure losses in the HTF generated across the thermal store.
In the case of a packed bed of particles, the relationship between fluid flow rate and pressure loss per unit length has been described by Ergun as:
                                          Δ            ⁢                                                  ⁢            P                    L                =                                            150              ⁢                                                          ⁢              μ              ⁢                                                          ⁢                                                u                  ⁡                                      (                                          1                      -                      ɛ                                        )                                                  2                                                                    D                p                2                            ⁢                              ɛ                3                                              +                                    1.75              ⁢                                                          ⁢              ρ              ⁢                                                          ⁢                                                u                  2                                ⁡                                  (                                      1                    -                    ɛ                                    )                                                                                    D                p                            ⁢                              ɛ                3                                                                        (        1        )            where:
ΔP/L is the pressure drop per unit length;
u is the fluid velocity;
μ is the fluid viscosity;
ε is the void space of the bed (i.e. the ratio of the volume of space unfilled by particles to the total volume of the bed;
Dp is the diameter (i.e. equivalent spherical diameter) of the particles; and
ρ is the fluid density.
Therefore, optimisation of the fluid velocity, particle diameter and shape of the particles is essential to minimise the pressure drop per unit length and, consequently, the HTF pumping losses.
A number of empirical relationships have been proposed to describe the heat transfer process between a fluid and particle bed by relating the Nusselt number (Nu), Reynolds number (Re), and Prandtl (Pr) number. For example, Ranz & Marshall proposed the following relationship:Nu=2+1.8(Re)0.5(Pr)0.33  (2)
The Reynolds number is defined as:
                    Re        =                              νρ            ⁢                                                  ⁢            l                    μ                                    (        3        )            and the Nusselt number is defined as:
                    Nu        =                  hl          k                                    (        4        )            
where v is the HTF velocity, ρ is the HTF density, μ is the HTF viscosity, h is the heat transfer coefficient between the HTF and the particles, k is the HTF conductivity and l is the relevant characteristic length. As the Prandtl number only relates to the physical properties of the HTF, it can be concluded that the heat transfer coefficient (h) is proportional to the square root of the HTF velocity (v)
Inspection of equations (1), (2), (3) and (4) indicates that pressure losses are proportional to the square of velocity, whereas convective heat transfer is proportional to the square root of velocity.
The present inventors have determined that careful optimisation of the flow rate through the packed bed is essential if the pressure drop is to be controlled within acceptable limits but good heat transfer between the HTF and the thermal mass is to be achieved.
For example, it is desirable to limit the pressure drop across a particular thermal store to 0.5 bar (50 kPa) as a vessel below that pressure is generally not classified as a pressure vessel and is therefore less expensive to manufacture. The inventors have determined that a Nusselt number of greater than 100 is preferable in order to ensure good heat transfer. FIG. 3 shows the predicted performance of such a thermal store across a range of flow rates. It can be seen that there is a narrow ‘operating window’ of HTF flow rates between about 1 kg/s and 2 kg/s where the store will operate within these specified limits. In a thermal energy storage system, it would be desirable to be able to discharge the store at a rate of about 20% of the charging rate. In that case, a fixed geometry store would suffer either poor thermal performance during discharging or unacceptably high pressure loss during charging.
The aspect ratio of a thermal mass is the ratio of the mean length of the thermal mass to the mean cross-sectional flow area. The inventors have determined that a small aspect ratio, i.e. a large flow area and/or a short length, is desirable for a given thermal mass to minimise the HTF velocity and therefore reduce pressure losses. However, such a small aspect ratio leads to high ‘end losses’ during the charging and discharging of the store: during charging, the thermal energy from the HTF cannot be completely captured unless an over-long store is used. This is undesirable as the final section of the store near the output of the store is not fully charged and thermal energy flows between the charged and partially charged sections during the storage phase of the cycle, resulting in a degradation of thermal efficiency. A similar problem is encountered during discharge; as the store discharges, the outlet temperature deviates from the storage temperature at the end of the cycle and it is not possible to fully discharge the final section of the store near the outlet without, again, resulting in a loss of thermal efficiency. This is illustrated in FIGS. 4 and 5 which show simulation results for thermal storage devices in which the charging flow is five times greater than the discharge flow. The left-most line in FIGS. 4 and 5 shows a plot of the temperature of the store at its inlet over time. The central line in FIGS. 4 and 5 shows a plot of the temperature of the store in its middle over time. The right-most line in FIGS. 4 and 5 shows a plot of the temperature of the store at its outlet over time. The area shaded in FIG. 4 represents the potential thermal losses due to end effects for the charging flow condition. The discharge process is stopped when the outlet temperature of the store, shown by the right most line in FIG. 5 is too high, leaving part of the store in a partially discharged state. The losses are about double for the lower flow rate case relative to the high flow rate case, as the store has been optimised for a higher flow rate.
Accordingly, there is a need for a thermal energy storage device and method which can be charged and discharged at different rates. There is also a need for a thermal energy storage device and method which can have a longer charge phase than discharge phase, or a longer discharge phase than charge phase.